Method to identify automatic transmission lubrication oil flow rates corresponding to a running vehicle without direct oil flow measurements

ABSTRACT

A method of determining automatic transmission lubrication fluid flow rates corresponding to a running vehicle without direct oil flow measurements is disclosed. A set of in-vehicle clutch torques for a chosen clutch pack during a gear shift event for a set of shift conditions is obtained. A series of bench tests at various clutch-pack clearances and oil-flow rates for the set of shift conditions are performed. The clearances and oil-flow rates are adjusted in response to the measured magnitudes exceeding thresholds. In-vehicle transmission lubrication oil-flow rates are estimated at the chosen clutch pack for the set of shift conditions when the bench-test and in-vehicle clutch torques are less than the thresholds. The steps are reproduced for other engine conditions and fluid temperatures corresponding to other transmission gear positions. A functional map of in-vehicle oil flow rates are produced, and the transmission is adjusted based on the map.

TECHNICAL FIELD

This invention relates generally to a bench test method for calibratingan oil-lubricated wet clutch for enabling an accurate replication ofdynamic engagement torque behaviors during a transmission shift event ina vehicle. More particularly, the invention relates to systematicallyadjusting and identifying initial clutch clearance and lubricationconditions during bench testing.

BACKGROUND

Many vehicles are used over a wide range of drive conditions, includingboth forward and reverse movement. The powertrain systems, in particularinternal-combustion engines, however, have desirable operatingconditions, including engine speed range, where they run mostefficiently. Consequently, automotive transmissions capable ofefficiently transmitting power at a variety of speed ratios arefrequently employed. Transmission speed ratio is the ratio of inputshaft speed to output shaft speed. When the vehicle is at low speed, thetransmission is usually operated at a high speed ratio such that itmultiplies the engine torque for improved acceleration. At high vehiclespeed, operating the transmission at a low speed ratio permits an enginespeed associated with quiet, fuel efficient cruising.

A common type of automatic transmission includes a gearbox capable ofalternately establishing a fixed number of power flow paths, eachassociated with a fixed speed ratio. The gearbox includes a number ofshift elements such as wet clutches and brakes, where their frictionalinterfaces are continually lubricated with automatic transmission fluid.A particular power flow path is established by engaging a particularsubset of the shift elements. To shift from one power flow path toanother power flow path with a different speed ratio, one or more shiftelements must be released while one or more other shift elements must beengaged. Some shift elements are passive devices such as one wayclutches, while other shift elements engage or disengage in response tocommands from a controller. For example, in many automatictransmissions, the shift devices are hydraulically controlled frictionelements such as wet clutches or brakes. The controller regulates thetorque capacity of the shift element by regulating an electrical currentto a solenoid, which adjusts a force on a valve which, in turn, adjustsa pressure in a hydraulic circuit.

A modern automatic transmission is controlled by a microprocessor whichadjusts the torque capacity of each wet shift element, including anylock-up clutch, at regular intervals. At each interval, the controllergathers information indicating the driver's intent, such as thepositions of the shifter (PRNDL), the accelerator pedal, and the brakepedal. The controller also gathers information about the currentoperating state of the vehicle, such as speed, and of the engine.Increasingly, information is also available from other sources, such asanti-lock brake controllers and GPS systems. Using this information, thecontroller determines whether to maintain the currently establishedpower flow path or to shift to a different power flow path. If thecontroller decides to shift to a different power flow path, thecontroller then adjusts the torque capacities of the off-going shiftelements and the on-coming shift elements in a coordinated manner inorder to make the transition as smooth as possible. However, it remainsa challenge to accurately deliver desired clutch torque for bothengagement and release processes. This is because the wet clutchhydrodynamically transmits torque by means of viscous shear across fluidfilm between rotating clutch plates with or without mechanical asperitycontact at the frictional interfaces. This hydrodynamic torque isparticularly sensitive to fluid conditions at the interface.Specifically, the amount of hydrodynamic torque is affected by thechange rate of oil film thickness and slip speed during clutchengagement and exhibits highly non-linear behaviors with respect toactuator force profile and slip speed, making it difficult for thecontroller to consistently deliver desired torque under all shiftconditions.

A wet clutch bench tester is widely utilized in order to improve clutchdesign features during a transmission development process. Theindustry-standard clutch test stand, which is often referred to as SAENo. 2 tester, is an inertia-absorption-type brake machine, typicallyequipped with a pneumatic actuator with limited control authority. It isutilized for evaluating clutch performance stability and durabilityduring engagement duty cycles, but not capable of recreating realisticclutch slip and actuator force profiles for the purpose of shift controldevelopment. There are other clutch testers with advanced features suchas enhanced electrical motor control and programmable hydraulicactuator, enabling the use of the methodology patented in U.S. Pat. No.6,923,049 for accurately replicating clutch slip and actuator forceprofiles during torque phase and inertia phase of shifting as observedin a vehicle. However, clutch torque measurements obtained from suchadvanced testers do not correlate well with those observed in a vehicle,even if slip and force profiles are accurately replicated. There is aneed to invent and establish a clutch bench test methodology thatenables accurate replication of clutch engagement torque behaviors, asobserved in a vehicle, to support transmission shift controldevelopment.

SUMMARY

According to one embodiment, a method of determining automatictransmission lubrication fluid flow rates corresponding to a runningvehicle without direct oil flow measurements is provided. A set ofin-vehicle clutch torques for a chosen clutch pack during a gear shiftevent for a set of shift conditions is obtained. A series of bench testsat various clutch-pack clearances and oil-flow rates for the set ofshift conditions are performed. The clearances and oil-flow rates areadjusted in response to the measured magnitudes exceeding thresholds.In-vehicle transmission lubrication oil-flow rates are estimated at thechosen clutch pack for the set of shift conditions when the bench-testand in-vehicle clutch torques are less than the thresholds. The stepsare reproduced for other engine conditions and fluid temperaturescorresponding to other transmission gear positions. A functional map ofin-vehicle oil flow rates are produced, and the transmission is adjustedbased on the map. The designs and controls of the transmission are thenadjusted based on the functional map of in-vehicle transmissionlubrication oil flow rates for desired lubrication oil distributionswithin the transmission.

According to another embodiment, a method includes obtaining in-vehicleclutch torques (IVCTs) for a clutch pack during a gear shift event in atransmission, then performing bench tests at various lubricationoil-flow rates to obtain bench-test measured clutch torques (BTMCTs),then adjusting the oil-flow rates during the performing based on adifference between the BTMCTs and the IVCTs, and then constructing thetransmission accounting for an estimated in-vehicle lubrication oil-flowrate that is based on the difference being less than a threshold.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic representation of a transmission, according to oneembodiment.

FIG. 2 is a graph illustrating the dynamic response of hydraulicpressure at the actuator piston to a shift element command signal,according to one embodiment.

FIG. 3 is a graph illustrating how the dynamic response of hydraulicpressure at the actuator piston to a shift element command signal mayvary based on environmental conditions, according to one embodiment.

FIG. 4 is a flow chart for controlling a shift element, according to oneembodiment.

FIG. 5 is a flow chart for controlling a shift element and adapting ashift element transfer function while accounting for the dynamicresponse, according to one embodiment.

FIG. 6 is a flow chart for controlling a shift element and adapting ashift element transfer function when desired clutch torque can beforecast in advance, according to one embodiment.

FIG. 7 is a flow chart of bench test procedure for calibrating a clutchpack clearance and an oil flow rate into a clutch based on a desiredclutch torque profile for a vehicle shift condition, according to oneembodiment.

FIG. 8 is a flow chart of bench test procedure for generating a clutchelement transfer function based on a set of calibrated clutch packclearances and oil flow rates, according to one embodiment.

FIG. 9 is a flow chart of a method for generating and mapping oil flowat a first clutch C1 in first gear, according to one embodiment,according to one embodiment.

FIG. 10 is an example of an oil flow map at clutch C1 at first gearcreated from the test procedures accomplished according to the method ofFIG. 9, according to one embodiment.

FIG. 11 is a flow chart of a method for generating and mapping oil flowat additional clutches Cn in a corresponding number n of differentgears, according to one embodiment.

FIG. 12 is an example of an oil flow map at a clutch Cn at an n-th gearcreated from the test procedures accomplished according to the method ofFIG. 11, according to one embodiment.

FIG. 13A illustrates an oil flow map at the first clutch C1 at firstgear with two different oil temperatures as engine speed varies; FIG.13B illustrates another oil flow map, but for a second clutch C2 at asecond gear; FIG. 13C illustrates another oil flow map for anotherclutch Cn for an n-th gear.

DETAILED DESCRIPTION

Embodiments of the present disclosure are described herein. It is to beunderstood, however, that the disclosed embodiments are merely examplesand other embodiments can take various and alternative forms. Thefigures are not necessarily to scale; some features could be exaggeratedor minimized to show details of particular components. Therefore,specific structural and functional details disclosed herein are not tobe interpreted as limiting, but merely as a representative basis forteaching one skilled in the art to variously employ the embodiments. Asthose of ordinary skill in the art will understand, various featuresillustrated and described with reference to any one of the figures canbe combined with features illustrated in one or more other figures toproduce embodiments that are not explicitly illustrated or described.The combinations of features illustrated provide representativeembodiments for typical applications. Various combinations andmodifications of the features consistent with the teachings of thisdisclosure, however, could be desired for particular applications orimplementations.

Controlling a hydraulically actuated automatic transmission requiresmanipulating a number of pressure commands to achieve a desired result.The desired result may be, for example, an upshift or downshift withparticular torque and speed characteristics as a function of time. Foran upshift, for example, the desired result may be a torque transferphase that takes a specified amount of time, followed by a specifiedspeed ratio vs. time profile during the inertia phase. In open loopcontrol, the controller uses a model of the transmission to calculatewhat pressure commands will produce the desired result and then commandsthose pressure values. The model may be an empirical model based ontesting a representative transmission or may be derived from physicallaws and nominal transmission characteristics such as dimension.However, the actual behavior of the transmission may differ from themodel for several reasons. First, there are part to part variationsamong transmissions of the same design. Second, a particulartransmission varies over time due to gradual wear or unusual events.Third, the transmission responds to a large number of environmentalfactors such as temperature, atmospheric pressure, etc.

To improve control in the presence of these variations, called noisefactors, a controller may utilize closed loop control. Closed loopcontrol improves the result within a particular event, such as a shift.In closed loop control, the controller measures the property thatdefines the desired behavior, such as speed ratio. The differencebetween the measured value and a target value is called the error. Thecommanded pressure is set to the open loop term plus one or more closedloop terms that are functions of the error. Widely used examples of suchfunction include linear terms such as: a proportional term (p term), aderivative term (d term), and an integral term (i term). Each suchlinear closed loop term has a coefficient of proportionality. Thesecoefficients are set during calibration such that, despite the presenceof noise factors, the result converges rapidly toward the desiredbehavior with minimal oscillation. Non-linear feedback terms may beemployed in order to account for changing operating conditions, or tocompensate for known non-linearities in the control system.

Adaptive control improves the result over a number of events. After anevent, the controller utilizes the measurements made during the event torevise the model. (Sometimes this is done implicitly rather thanexplicitly, such as by modifying the open loop terms.) As the modelbecomes more representative of the particular transmission and thepresent conditions, the open loop control of future events becomesbetter. This minimizes the error that the closed loop terms need toaccommodate. Moreover, it improves robustness of the phases of the shiftthat lack feedback information (e.g., the torque-transfer phase).

Both closed loop control and adaptive control require measurement orestimation of the properties that define the desired behavior. Ideally,this would be accomplished by having a separate sensor for eachproperty. Unfortunately, sensors add cost and weight to a design andintroduce failure modes. Also, some parameters are difficult to measurebecause the sensor would need to be buried in an inaccessible locationof the transmission. Consequently, in practice, the number and type ofsensors is restricted. When there is no sensor for the property thatdefines the desired behavior, a model may be utilized to estimate thevalue based on the available measured properties. These models aresubject to the same types of noise factors as the models used to computethe open loop terms. Furthermore, a model may include assumptions thatmake it valid only under certain operating conditions, such as when in2nd gear. In order to estimate the property in all of the relevantoperating conditions, the controller may need to use multiple models. Insome operating conditions, more than one of the models may be valid,leading to possibly conflicting estimates. In such cases, the controllermust determine which estimate to trust. The controller may use thetrusted model to revise the other models in order to improve theestimate in operating conditions in which the trusted model is unusable.

FIG. 1 illustrates a representative front wheel drive automatictransmission. The transmission is contained in a housing 10 that isfixed to vehicle structure. An input shaft 12 is driven by the vehicleengine. The input shaft may be connected to the engine via a damper thatisolates the transmission from engine torque pulsations. An outputelement 14 drives vehicle wheels. The output element 14 may be driveablyconnected to the wheels via final drive gearing and a differential. Thefinal drive gearing transmits the power to a parallel axis andmultiplies the torque by a fixed final drive ratio. The final drivegearing may include layshaft gears, a chain and sprockets, and/orplanetary gearing. The differential divides the power between left andright front wheels while permitting slight speed differences as thevehicle turns. Some vehicles may include a power take-off unit thattransfers power to rear wheels.

A torque converter 16 has an impeller 18 fixed to input shaft 12 and aturbine 20 fixed to turbine shaft 22. Torque converter 16 transmitstorque from input shaft 12 to turbine shaft 22 while permitting turbineshaft 22 to rotate slower than input shaft 12. When turbine shaft 22rotates substantially slower than input shaft 12, a torque converterstator 24 is held against rotation by one way clutch 26 such that thetorque applied to turbine shaft 22 is a multiple of the torque suppliedat input shaft 12. When the speed of turbine shaft 22 approaches thespeed of input shaft 12, one way clutch 26 overruns. Torque converter 16also includes a lock-up clutch 28 that selectively couples input shaft12 to turbine shaft 22.

Gear box 30 establishes a number of speed ratios between turbine shaft22 and output element 14. Specifically, gear box 30 has three planetarygear sets and five shift elements that establish six forward and onereverse speed ratio. Simple planetary gear sets 40, 50, and 60 each havea sun gear (42, 52, 62), a carrier (44, 54, 64), and a ring gear (46,56, 66) that rotate about a common axis. Each planetary gear set alsoincludes a number of planet gears (48, 58, 68) that rotate with respectto the carrier and mesh with both the sun gear and the ring gear.Carrier 44 is fixedly coupled to ring gear 66 and output element 14,carrier 54 is fixedly coupled to ring gear 46, ring gear 46 is fixedlycoupled to carrier 64, and sun gear 52 is fixedly coupled to turbineshaft 22.

The various speed ratios are established by engaging variouscombinations of shift elements. A shift element that selectively holds agear element against rotation may be called a brake whereas a shiftelement that selectively couples two rotating elements to one anothermay be called a clutch. Clutches 72 and 74 selectively couple turbineshaft 22 to carrier 64 and sun gear 62, respectively. Brakes 76 and 78selectively hold sun gear 62 and sun gear 42, respectively, againstrotation. Brake 80 selectively holds carrier 64 against rotation.Finally, one way clutch 82 passively holds carrier 64 against rotationin one direction while allowing rotation in the opposite direction.Table 1 illustrates which shift elements are engaged to establish eachspeed ratio.

TABLE 1 72 74 76 78 80/82 Ratio Step Reverse X X −3.00 71% 1st X X 4.202nd X X 2.70 1.56 3rd X X 1.80 1.50 4th X X 1.40 1.29 5th X X 1.00 1.406th X X 0.75 1.33

Shift element or wet friction elements 72-80 may be hydraulicallyactuated multi-plate wet friction clutches or brakes. Controller 84controls the pressure of transmission fluid routed to each shift elementactuator. This controller may adjust an electrical current to one ormore variable force solenoids to control the pressure supplied to eachclutch actuator piston. When pressurized fluid is first supplied to ashift element actuator, it moves a piston into a stroked position. Then,the piston forces plates together causing the shift element to transmittorque. Shift elements such as wet clutches and brakes are continuallylubricated with transmission fluid at their frictional interfaces.Accordingly, the wet clutch hydrodynamically transmits drag torquebetween rotating plates even if actuator piston is retracted. However,the torque capacity is considered practically negligible for shiftcontrol purpose until the piston reaches the so-called touch point whichis subject to unit-to-unit hardware variability, component wear andenvironmental conditions such as transmission fluid temperature. Moreprecisely, as the piston is stroked, lubrication film at the frictionalinterface is squeezed out. When the film thickness becomes sufficientlysmall, the clutch can hydrodynamically transmit a significant amount oftorque through viscous shear of rotating plates even before asperitycontact takes place between rotating plates. This hydrodynamic torque ishighly sensitive and non-linear with respect to the rate of filmthickness change which is affected by actuator force profile and fluidconditions such as temperature, making it difficult to predict andcontrol the touch point where the clutch starts transmitting usabletorque for the purpose of shifting. Once oil film is sufficientlysqueezed out, surface asperity contact takes place, enabling torquetransfer through mechanical friction. As the engagement processcontinues, hydrodynamic torque diminishes and mechanical friction torqueeventually constitutes the entire clutch torque. Replicating andpredicting this complex dynamic engagement process on a clutch teststand, accurately capturing a fine balance of hydrodynamic torque andmechanical friction torque, is considered particularly difficult. Whenthe pressure is relieved, a return spring moves the piston to a released(not stroked) position, freeing a torque path. The controller receivessignals from transmission sensors such as a turbine speed sensor 86 andan output speed sensor 88. A typical upshift includes three phases: apreparatory phase, a torque transfer phase, and an inertia phase. Duringthe preparatory phase, pressure is commanded to the on-coming shiftelement in order to stroke the piston so that it is ready forengagement. Also, the torque capacity of the off-going shift element maybe reduced from a holding capacity well in excess of the transmittedtorque to a value close to the actual transmitted torque. During thetorque transfer phase, the torque capacity of the off-going shiftelement may be gradually reduced while the torque capacity of theon-coming shift element is gradually increased. During this phase, thereis little or no slip across the off-going shift element but considerableslip across the on-coming shift element. When the off-going shiftelement torque capacity becomes sufficiently small, allowing it to slip,the power flow path associated with the upshifted gear is establishedthrough rising on-coming clutch engagement toque. Therefore, the torqueratio is equal to the upshifted torque ratio. However, the speed ratiois still equal or nearly equal to the original speed ratio. When theoff-going shift element is completely released, the torque transferphase ends and the inertia phase begins. During the inertia phase, thetorque capacity of the on-coming shift element is controlled toeliminate the slip across the on-coming shift element and bring thespeed ratio to the upshifted speed ratio in a controlled manner.

A power-on downshift also includes an inertia phase and a torquetransfer phase, although they occur in the opposite order. During theinertia phase, the torque capacity of the off-going shift element iscontrolled to bring the speed ratio to the downshifted speed ratio in acontrolled manner, which involves a progressively increasing slip acrossthe off-going shift element. The on-coming shift element may be preparedfor engagement by commanding pressure in order to stroke the piston.During the torque transfer phase, which occurs after the inertia phase,the torque capacity of the previously stroked on-coming shift element isgradually increased while the torque capacity of the off-going elementis reduced to zero.

During the shift, accurate control of torque capacity is important inorder to achieve a smooth shift. For example, during the torque transferphase, the increase in torque capacity of the on-coming shift elementmust be carefully coordinated with the decrease in torque capacity ofthe off-going shift element. If the torque capacity of the on-comingshift element is ramped up too slowly, relative to the input torque andthe rate of decrease of off-going shift element torque capacity, then anengine flare occurs. If, on the other hand, the on-coming shift elementtorque is ramped up too quickly, then a tie-up condition occurs. Bothresult in an excessive decrease in output torque.

Open loop control of shifts is aided by having a model for each shiftelement. The torque capacity of each clutch is adjusted by adjusting anelectrical current to a solenoid in the hydraulic control system valvebody. A valve in the valve body responds by adjusting the pressure in afluid circuit in proportion to the force generated by the solenoid. Thecontrol fluid is routed to a clutch apply chamber where it pushes apiston to compress a clutch pack with interleaved friction plates andseparator plates, squeezing out lubrication oil film at the clutchfrictional interfaces, developing hydrodynamic torque before mechanicalfriction torque starts rising. A return spring forces the piston backwhen the pressure is relieved. In an exemplary steady state model of ahydraulically actuated friction clutch or brake, the torque capacity isa function of the electrical current supplied. This function generallyhas two segments. In a first segment, from zero current up to thecurrent required to overcome the force of the return spring, the torquecapacity is assumed to be practically zero. Beyond the current requiredto overcome the return spring, the torque capacity is assumed toincrease linearly with respect to the current. However, the steady statemodel does not and cannot capture actual clutch behaviors, in particularat low temperatures where clutch torque is highly non-linear due tohydrodynamic torque and sensitive to time-dependent actuator forceprofiles. In an alternative model, the fluid pressure is a function ofthe electrical current and the torque capacity is a function of thefluid pressure. This alternative model may be useful if a pressuresensor is available to provide a pressure feedback signal. In somemodels, other factors such as temperature may be considered. The shiftelement model is represented by a transfer functionT _(cl) =F(U,X)where T_(cl) is the predicted clutch torque, U is the command signal,such as current or pressure, and X is a set of parameters indicating theenvironmental conditions, such as temperature.

In addition to consideration of the steady state relationship betweenclutch torque and a command signal, such as a pressure, the model mayconsider dynamic effects of hydraulic actuator system, accounting forpressurized fluid movement from the valve body to the clutch piston.FIG. 2 represents a possible model of the dynamic response of hydraulicpressure at the actuator piston with respect to a control signal. Inthis example, the commanded control signal 100 changes from one level toanother level in a step function. The actual control signal 102 does notimmediately change to the second level. Instead, the actual controlsignal remains at the original level for a period of time called thepure delay τ_(d). Then, the actual control signal asymptoticallyapproaches the second according to a first order distributed delay witha time constant of τ. After a delay of τ, the actual signal has changed63.2% of the way to the second value. This dynamic response model may berepresented by the dynamic transfer function

${G(s)} = {e^{{- \tau_{d}}s}\frac{1}{{\tau\; s} + 1}}$

As illustrated by FIG. 3, the dynamic response of hydraulic pressure atthe actuator piston with respect to a commanded control signal may varydepending upon environmental conditions such as fluid temperature withinthe hydraulic actuator system. For example, curve 102 represents thedynamic behavior in one environmental condition X₁ while curve 104represents the dynamic behavior in a second environmental condition X₂.For example, X₁ may correspond to normal operating temperature and X₂may correspond to a colder temperature. The impact of environmentalconditions may be modeled by expressing the model parameters τ_(d) and τas functions of a set of environmental condition parameters X. Theactuator models described in FIGS. 2 and 3 are intended for capturingdynamic behaviors of pressurized fluid in the hydraulic control system,which may be observed using pressure transducers instrumented in thevalve body and clutch piston cylinder. However, they do not and cannotdescribe non-linear hydrodynamic behaviors of oil film that transmitshydrodynamic torque during clutch engagement process. Accordingly, evenif the actuator dynamics is accounted through the actuator models asillustrated in FIG. 2 and FIG. 3, the model does not and cannot captureactual clutch behaviors, in particular at low temperatures wherehydrodynamic clutch torque is highly non-linear and sensitive withrespect to slip and actuator force profiles. There are no known dynamicclutch engagement models that can robustly determine clutch torques,accounting for both hydrodynamic torque and asperity contact torque,under all operating conditions, in particular between −40 C and +30 C.

Several of the models described above can be represented in controller84 as one or more lookup tables. A lookup table stores predicted valuesof a model output variable for various combinations of values of one ormore model input variables. When there is only one input variable, thelookup table is referred to as one dimensional. For example, a onedimensional lookup table may be used to represent the clutch transferfunction model by storing values of clutch torque capacity at variouscommanded pressures. When the output variable is dependent upon multipleinput variables, higher dimensional lookup tables are used. For example,a clutch transfer function may be represented as a two dimensionallookup table based on pressure and temperature.

To find a value for a model output variable based on particular valuesof the model input variables, the controller finds the stored pointsthat are closest to the particular values and then interpolates. To findan input variable corresponding to a desired output variable, reverseinterpolation is used. This reverse interpolation yields a uniquesolution only when the underlying function is monotonic. Alternatively,the model may be re-formulated such that clutch torque is an inputvariable and commanded pressure is an output variable. In practice,there are no known models that can accurately determine dynamic clutchengagement torques, accounting for both hydrodynamic torque and asperitycontact torque, under all operating conditions in response totime-dependent slip and actuator force profiles. Thus, the reserveinterpolation based on such models often fails to generate correctcommand pressures, resulting in poor shift quality.

Several methods are known for adaptively updating a model represented asa lookup function. These include both stochastic adaptation methods andperiodic adaptation methods. Stochastic adaptation methods update thevalues in the lookup table in response to individual observed results.One such method is described in European Patent Application EP 1 712 767A1, which is incorporated by reference herein. When the observed resultdiffers from the value estimated by the lookup table, the stored valuesfor nearby values of the model input variables are modified such that anew prediction for the same model input values is closer to the observedresult. For stability, the adaptation is not allowed to change thestored values by too much at once. The adaptation may be restricted invarious ways. For example, adaptation may only be allowed when theoperating point is sufficiently close to one of the stored values. Also,there may be pre-defined bounds outside which adaptation is notperformed. In a periodic adaptation method, multiple observations arestored and then a curve fitting process is performed to calculate newvalues for model parameters. As with stochastic adaptation methods,there may be restrictions on the rate of adaptation and there may beboundaries beyond which adaptation is not permitted. In practice, theeffectiveness of adaptation strategies is limited to the inertia phaseof shifting, where clutch torque behaviors are readily observablethrough component speed changes using commonly available sensors. Thereis a need for robust adaptation methods for capturing highly non-linearclutch torque behaviors during torque transfer phase of shifting, whereclutch torque observability is limited.

In practice a transmission shift control calibration may be accomplishedbased on the combined use of steady-state clutch torque model, dynamichydraulic actuator model, adaptive strategies as well as closed-loopcontrols, to manage and control clutch behaviors for shift conditionswhere clutch behaviors are observable. Open loop clutch controls may beutilized together with steady-state-clutch model and dynamic hydraulicactuator model for shift conditions where clutch torque is notobservable. The models may be derived theoretically or empiricallythrough clutch bench testing. However, the unavailability of accurateclutch torque models in particular at low temperature requires atransmission engineer to manually tune shift control parameters based ontime-consuming trial- and error approaches in a vehicle. Thus, amethodology is desired to systematically capture and accuratelyrepresent dynamic clutch engagement behaviors over desired shiftconditions and to construct a dynamic clutch transfer function whichaccurately describes a relationship between clutch torque and actuatorforce.

FIG. 4 illustrates a clutch control algorithm that utilizes a statictorque capacity model of the clutch system for control. Solid linesindicate flow of control. Dotted lines indicate flow of information. At110, the controller determines the operating conditions X. At 112, thecontroller determines the desired clutch torque T_(des) which is equalto a function T_(ctrl). T_(ctrl) may be based on indicators of driverintention such as accelerator pedal position, on estimates ormeasurements of transmission input torque, and on measurements from thetransmission system, such as the speeds of various elements. Forexample, during the inertia phase on an upshift, the information fromspeed sensors may be used to determine how quickly the shift isprogressing. If the shift is progressing more slowly than desired,T_(ctrl) may be increased. At 114, the commanded control signal U_(com)is computed using an inverse of the shift element transfer functionwhich provides a relationship between commanded signal and clutchtorque. At 116, the controller issues the computed control signal to theactuators. At 118, the controller determines if the shift event hascompleted and repeats the process if it has not. The performance of theshift control algorithm illustrated in FIG. 4 is limited by the staticclutch transfer function which typically assumes a Coulomb frictionmodel without accounting for hydrodynamic torque transfer mechanisms.The static transfer function may be based on clutch test data from SAENo2. Clutch test stand whose correlation with vehicle shift data isknown to be less than satisfactory for most shift conditions. Theperformance of the shift control algorithm in FIG. 4 can besubstantially improved if a dynamic clutch transfer function is madeavailable, capable of predicting in-vehicle clutch behaviors under allshift conditions.

The algorithm of FIG. 4 can be improved by adapting the static transferfunction using a measured clutch torque if such measurements areavailable. However, due to the dynamic response of the hydraulic clutchactuator system as illustrated in FIGS. 2 and 3, one would not expectthe static transfer function alone, even if it is adapted with measuredclutch torque, to accurately relate the present command control signalto the present measured torque when the control signal is changing. Theclutch control algorithm of FIG. 5 utilizes the dynamic transferfunction of the hydraulic clutch actuator to account for the dynamicresponse of the actuator system while adapting the clutch transferfunction. At 120, the controller estimates the actual effective controlsignal, U_(act), using the dynamic transfer function of the actuator anda recorded profile of past commanded control signals. At 122, the actualeffective control signal is used with the static torque transferfunction of the clutch to predict the present clutch torque T_(cl). At124, the controller estimates the present clutch torque based onmeasurements. Methods for doing this are described in U.S. Pat. No.8,510,003 and U.S. patent application Ser. No. 14/668,062 now U.S. Pat.No. 9,709,164 which are hereby incorporated by reference herein. Thesetwo clutch torque estimates are compared at 126 to compute an errorterm. At 128, the static clutch transfer function is adapted to reducethe error. Since the clutch transfer function is adapted only a smallamount during each iteration, random noise in the measurements does notcause substantial adaptation. This adapted clutch transfer function isused at step 114′ to more accurately compute the control signal.Optionally, the desired clutch torque T_(des) may include a function ofthe error at 112′.

The algorithm of FIG. 5 also utilizes a dynamic model of the hydraulicclutch actuator to compute the commanded control signal U_(com). Insteadof computing U_(com) directly from T_(des) in a single step as 114 ofFIG. 4, the calculation is divided into two steps 114′ and 130. At 114′the controller computes the desired control signal U_(des) using thestatic clutch transfer function. Then, at 130, the controller uses alead-lag filter to at least partially compensate for the dynamicsresponse of hydraulic actuator system. Ideally, the controller would usethe inverse of the dynamic response function G⁻¹( ) of the hydraulicactuator system. However, the dynamic response function of the hydraulicactuator system may not be invertible without information about futurevalues of U_(des). Consequently, it may be necessary to use a lead-lagfilter with a transfer function

${G^{{- 1}*}(s)} = \frac{{\tau_{1}s} + 1}{{\tau_{2}s} + 1}$that approximates G⁻¹. The lead time constant τ₁ may be selected equalto the first order time constant τ. Alternatively, to also partiallycompensate for the pure time delay, τ₁ may be selected equal to the sumof the first order time constant and the pure delay τ+τ_(D). The lagtime constant τ₂ is selected such that τ₂ is much smaller than τ₁ butstill large enough to prevent excessive sensitivity to small variationsin U_(des).

FIG. 6 illustrates a further improvement upon the clutch controlalgorithm of FIG. 5. At 112″, the controller predicts not only thepresent desired clutch torque, but predicts the desired clutch torqueover a period of time extending into the future. This is possiblebecause some of the terms used to compute the desired clutch torque areknowable or predictable in advance. For example, the nominal(feedforward) may be known if the desired ratio change is known for thenext several update loops. Also, a feedback term based on an integral ofthe error may be predicted by assuming that the error continues at thepresent level. The controller may use the present value for other terms.Specifically, the desired clutch torque is predicted for a period oftime at least as long as the pure time delay τ_(d). Then, at 114″, thedesired control signal is computed based on the static clutch torquetransfer function and the predicted desired clutch torque for the sametime period. At 130′, the commanded control signal is computed byapplying the lead-lag filter to the predicted desired control signalτ_(d) in the future. Consequently, the control signal has time to takeeffect by the time that clutch torque is actually desired despite thedelays due to the hydraulic actuator system dynamics. The shift controlalgorithms illustrated in FIGS. 5 and 6 require accurate clutch torquemeasurements or estimates in real-time for the entire shift duration,including torque phase and inertia phase. In-vehicle clutch torque maybe available only in a test vehicle which may be equipped withtransmission torque sensor. Accurate measurements or estimates ofin-vehicle clutch torque is not generally available in line-producedvehicles, making it difficult to directly employ the algorithmsillustrated in FIGS. 5 and 6. In the absence of real-time clutch torquemeasurements in volume production vehicles, it become more important todevelop a predictive dynamic clutch torque transfer function throughbench tests to allow a shift controller to accurately determine clutchtorque in response to control signals.

Step-ratio automatic transmissions require a complex sequence of wetclutch controls in order to shift between available gear positions. Aconventional clutch control involves manual adjustments of shiftcalibration parameters, such as clutch actuator force profile for eachshift, that are stored as a multi-dimensional look-up table inpowertrain control strategies on a powertrain control module. Suchapproach is becoming increasingly undesirable because of a large numberof shift combinations between gear positions.

An alternative approach is to utilize a functional representation ofclutch behaviors, often referred to as clutch transfer functions (CTF),which may require less number of shift calibration parameters ascompared to manual adjustments of actuator force profiles. CTFrepresents a relationship between clutch actuator force (or pressure)and torque transmitted through the clutch pack. CTF may take a differentform such as a multi-variable lookup-table, polynomial function orneural net. A challenge is that there is no well-established methodologyto analytically or experimentally generate CTF that accuratelyrepresents realistic in-vehicle clutch engagement characteristics duringboth torque transfer phase and inertia phase of shifting for a broadrange of vehicle operating conditions.

A wet clutch pack is lubricated with transmission fluid at frictionalinterfaces. Torque transmission characteristics are highly sensitive toengagement conditions, specifically during torque transfer phase ofshifting where hydrodynamic torque is significant. A transmissioncontrol may assume a linear and static relationship between torque andactuator force for CTF. However, this approach may not account fornon-linear, dynamic nature of hydrodynamic clutch behaviors, which occurat specific phases of clutch engagement or disengagement, for example,during the torque phase at lower oil temperatures. A failure to utilizeaccurate CTF in clutch controls may result in undesirable transmissionbehaviors in a vehicle.

Several methods are known that enable calculation of dynamic(time-dependent) clutch torque during a shift event in a vehicle, basedon measurements that are commonly available in test vehicles. Two suchmethods are disclosed in U.S. Pat. App. Pub. No. 2014/0324308 and U.S.application Ser. No. 14/668,190 now U.S. Pat. No. 9,726,280, which areboth incorporated by reference herein. These methods may be utilized toidentify CTF in a vehicle under dynamically-changing conditions. SuchCTF may be referred to as dynamic clutch transfer functions (DCTF).However, it may be impractical to systematically cover the entire clutchoperating conditions through vehicle testing. Also, it may be difficultto repeat clutch engagements under the same condition in a vehicle forrepeatability verification because of uncontrolled factors such astransmission oil flow rate and temperature. U.S. Pat. No. 6,923,049,which is incorporated by reference herein, describes a clutch bench testmethodology that enables accurate replication of torque phase andinertia phase conditions that are consistent with in-vehicle operations.Clutch torque behaviors obtained from bench testing, however, may stilldiffer from actual in-vehicle clutch behaviors.

A methodology to replicate in-vehicle clutch behaviors through benchtesting to enable an efficient and systematic generation of a DCTF undera broad range of clutch operating conditions is disclosed herein. Asindicated above, U.S. Pat. No. 6,923,049 describes clutch bench testmethodology that enables accurate replication of torque phase andinertia phase conditions that are consistent with in-vehicle operations.More specifically, U.S. Pat. No. 6,923,049 recreates dynamic profiles ofclutch slip and the force of the clutch actuator that are observedduring the torque phase and the inertia phase of shifting in an actualvehicle. The input torque to a clutch pack is controlled to achieve atarget slip profile, while accounting for the effects of engine torquemodulation that is commonly utilized during inertia phase of shifting.However, the clutch torque behaviors obtained from U.S. Pat. No.6,923,049 may still differ from actual in-vehicle clutch behaviors. Amethod to calibrate clutch bench test conditions, based on the methodfrom U.S. Pat. No. 6,923,049, to match in-vehicle clutch behaviors and amethod to generate a DCTF over a broad range of operating conditions aredisclosed herein.

The clutch pack clearance or clutch plate clearance is the accumulatedtravel distance of all the clutch plates during clutch engagement.Clutch torque behaviors from bench tests may differ from in-vehicleclutch behaviors due to limited knowledge or unavailability of clutchpack clearance or piston position at the beginning of torque transferphase, which may be interchangeably referred to as the touch point,after piston stroking is initiated. At the touch point, the wet clutchstarts to hydrodynamically generate torque before the piston is fullystroked. The touch point is highly sensitive to shift conditions becausehydrodynamic torque depends on various factors such as fluidtemperature, slip velocity, and oil film thickness. Clutch torquebehaviors from bench tests may also differ from in-vehicle clutchbehaviors due to limited knowledge or unavailability of the lubricationoil flow rate into the a hydraulically operated clutch pack during ashift event. A transmission oil temperature is typically measured at atransmission sump in a vehicle and can be utilized to specify oiltemperature in clutch bench testing. It should be noted that thetransmission oil is the same as the lubrication oil that operates thevarious clutches within a transmission. A thermo-couple may be insertedinto the clutch plates of a clutch pack to measure fluid temperature atthe interface between clutch plates in a test vehicle to define benchtest conditions. However, there is no practical methodology to measurethe amount of actual lubrication oil flow that travels through clutchpack during a shift event in a vehicle. Therefore, a nominal flow rateis typically prescribed in bench tests which may contribute toinaccurate test results. It may also be difficult to identifytouch-point clutch pack clearance at the beginning of the torque phasewhen the clutch torque starts rising during shifting. The difference inclutch return spring mechanisms between bench test set up and in-vehicleclutch pack assembly introduces additional complexity to estimating theinitial clutch pack clearance. Therefore, a nominal clutch packclearance is typically prescribed in a bench test which may alsocontribute to inaccurate results. A systematic method to calibrate theinitial clutch pack clearance and lubrication flow rate in clutch benchtests to match in-vehicle clutch behaviors during both torque phase andinertia phase of shifting is described herein.

Referring to FIG. 7, a flow chart of a method 700 for calibrating aclutch pack clearance and an oil flow rate into a clutch based a desiredclutch torque profile for a vehicle shift condition is illustrated. Themethod is initiated at the start block 702. Next, the method moves on tostep 704 where an in-vehicle clutch engagement torque as a function oftime is determined during a shift event for selected shift conditionC_(i). At step 704, the in-vehicle clutch engagement torque may bemeasured or determined according to the methodology described in U.S.Pat. App. Pub. No. 2014/0324308 and/or U.S. application Ser. No.14/668,190. The in-vehicle clutch engagement torque T_(v)(t, C_(i)) forthe selected shift condition C_(i) is then recorded at step 704 as afunction of time t and the shift conditions C_(i) which include clutchoperating variables such as the lubrication oil temperature and controlvariables such as the time-dependent actuator force profile and the slipprofile of the clutch.

After step 704, the method moves on to step 706 where an initial clutchpack clearance D(C_(i)) for the selected shift condition C_(i) isassumed. Next, the method moves on to step 708 where an initiallubrication oil flow rate into the clutch Q(C_(i)) for the selectedshift condition C_(i) is assumed. Once the initial clutch pack clearanceD(C_(i)) and the initial lubrication oil flow rate Q(C_(i)) have beenassumed, the method moves on to step 710.

At step 710, a clutch engagement bench test is conducted to determine abench test measured clutch torque T_(b)(t, C_(i)) as a function of timet. The bench test may be conducted according to the methodologydescribed in U.S. Pat. No. 6,923,049, replicating torque phase andinertia phase conditions that are consistent with the selected shiftcondition C_(i) (including the slip profile of the clutch, the actuatorforce profile, and the lubrication oil temperature) together with theassumed initial clutch pack clearance D(C_(i)) and the assumed initialoil flow rate Q(C_(i)). The touch-point clearance D(Ci) may be adjustedby using shims inserted into the tested clutch pack and alsore-positioning a clutch actuator piston.

After step 710, the method moves on to step 712, where the clutch torqueprofile from bench test, i.e., the bench test measured clutch torqueT_(b)(t, C_(i)) for the selected shift condition is compared to thein-vehicle torque profile T_(v)(t, C_(i)), i.e., the in-vehicle clutchengagement torque for the selected shift condition C_(i). If thedifference between in-vehicle clutch engagement torque T_(v)(t, C_(i))and the bench test measured clutch torque T_(b)(t, C_(i)) is not lessthan a predetermined value DT, the method moves on to step 714 where theinitial clutch pack clearance D(C_(i)) and the initial lubrication oilflow rate Q(C_(i)) are systematically adjusted, unlike prior art clutchtest methodologies where initial pack clearance and oil flow rate areneither recognized nor adjusted exclusively for the purpose of achievinga target torque profile obtained from the vehicle. The method thenreturns to step 710. The loop consisting of steps 710, 712, and 714 isrepeated until the difference between in-vehicle clutch engagementtorque T_(v)(t, C_(i)) and the bench test measured clutch torqueT_(b)(t, C_(i)) is less than the predetermined value DT (i.e., the benchtest measured clutch torque profile becomes within the pre-determinedenvelope of in-vehicle clutch torque profile).

It should be noted that alternatively at step 712 it may be determinedwhether the difference between in-vehicle clutch engagement torqueT_(v)(t, C_(i)) and the bench test measured clutch torque T_(b)(t,C_(i)) is greater than a threshold. This alternative embodiment of step712 would function in a similar manner as the embodiment describedabove, however the “no” and “yes” determination coming out of the boxrepresenting step 712 would be reversed.

Once the difference between in-vehicle clutch engagement torque T_(v)(t,C_(i)) and the bench test measured clutch torque T_(b)(t, C_(i)) for theselected shift condition is less than the predetermined value DT themethod moves on to step 716, where an adjusted or calibrated initialclutch pack clearance D(C_(i)) and an adjusted or calibrated initiallubrication oil flow rate Q(C_(i)) are recorded. The adjusted orcalibrated initial clutch pack clearance D(C_(i)) and adjusted orcalibrated initial lubrication oil flow rate Q(C_(i)) are the valuesused during the bench test in step 710 that resulted in the differencebetween in-vehicle clutch engagement torque T_(v)(t, C_(i)) and thebench test measured clutch torque T_(b)(t, C_(i)) being less than thepredetermined value DT.

Once the adjusted or calibrated initial clutch pack clearance D(C_(i))and the adjusted or calibrated initial lubrication oil flow rateQ(C_(i)) are recorded, the method moves on to step 718. At step 718, therelationship between the bench test measured torque T_(b)(t, C_(i)) andan actuator force profile F(t, C_(i)), which is a function of time forthe selected shift condition C_(i), is recorded as a coordinate locatedon a first dynamic clutch transfer function DCTF1(C_(i)) or simply apart of DCTF(C_(i)) for the selected shift condition C_(i), relative tothe adjusted or calibrated initial clutch pack clearance D(C_(i)) andthe adjusted or calibrated initial lubrication oil flow rate Q(C_(i))into the clutch. The method then ends at step 720.

Referring to FIG. 8, a flow chart of a method 800 for generating aclutch element transfer function for additional shift conditions basedon a set of initially calibrated clutch pack clearances D(C_(i . . . N))and a set of initial oil flow rates Q(C_(i . . . N)) is illustrated. Themethod is initiated at the start block 802. Next, the method moves on tostep 804 where the method described in FIG. 7 is repeated for N vehicleoperating or selected shift conditions C_(i), where, i=1, 2, . . . N, togenerate the first dynamic clutch transfer function DCTF1.

Next, the method moves on to step 806 where a map of calibrated initialclutch pack clearances D(C_(i)) and oil flow rates Q(C_(i)), with i=1, .. . N, is generated. The map may be a lookup table or in a functionalform using a conventional regression method. The generated map allowsinterpolation or extrapolation of initial clutch pack clearances D andoil flow rates Q over a broad range of operating conditions C, which arenot limited to C_(i), (i=1, . . . N). Considering that a rotationalspeed of clutch plate and oil temperature have an impact on thelubrication oil flow rate, the initial clutch slip and oil temperaturemay be used as primary factors for a functional map of D(C) and Q(C),where D(C)=function(initial slip, oil temperature) andQ(C)=function(initial slip, oil temperature).

Once the functional map of calibrated initial clutch pack clearancesD(C) and oil flow rates Q(C) is generated at step 806, the method moveson to step 808. At step 808, a set of M shift conditions Ctf_(k), wherek=1, . . . M, are selected or constructed as desired without necessarilyrequiring any in-vehicle measurements and without specifically requiringin-vehicle clutch torque. M clutch engagement bench tests are conductedunder the conditions Ctf_(k), where k=1, . . . M, for the desired torquephase and inertia phase conditions of a clutch, including various clutchslip profiles, actuator force profiles, lubrication oil temperaturevalues. The bench test may be conducted according to the methodologydescribed in U.S. Pat. No. 6,923,049, replicating both torque phase andinertia phase of shifting. At step 808 lubrication oil flow ratesQ(Ctf_(k)) and clutch pack clearances D(Ctf_(k)) are specified for thebench test based on the functional representation of D(C) and Q(C) whichare derived at 806.

After step 808, the method moves on to step 810, where the relationshipsbetween bench test measured torque profile T_(b)(t, Ctf_(k)) and theactuator force profile F(t, Ctf_(k)) are recorded from the M bench tests(k=1, . . . , M) as the second dynamic clutch transfer function DCTF2.Then method then ends at step 812. Note that in-vehicle clutch torque isnot required for shift conditions Ctf_(k) (k=1, . . . , M), whereasin-vehicle clutch torque is required for C_(i) (i=1, . . . , N). Themethodology in the flow chart 800 enables the prediction of in-vehicleclutch torque T_(b)(t, Ctf_(k)) for the conditions Ctf_(k) through theuse of the functional map of D and Q which are systematically derivedbased on in-vehicle clutch torque T_(v)(t, C_(i)).

Referring to FIG. 9, a flow chart 900 of a method for generating andmapping oil flow at a first clutch C1 in first gear is illustrated. Atstep 902, a first value of i is set, i=1. At 904, shift data during gearshifting between first gear and second gear is collected from the testmethods described above. At step 908, an oil flow rate Q_(1st,C1) isidentified at the clutch C1 for a given engine speed and oil temperatureaccording to the method shown in FIG. 7 and described above. Because themethod in FIG. 7 for identifying the lubrication fluid flow rate islargely affected by the torque phase of shifting, the resultingQ_(1st,C1) represents the lubrication oil flow at C1 corresponding tothe first gear condition before the engine speed changes to the secondgear level. At step 910, the value of i is determined to see if it isequal to M, where M is a predetermined number. If i is not equal to M,then at 912 certain engine operating characteristics are changed, suchas the engine speed, which changes the conditions of the automatictransmission fluid (ATF). In other words, engine speed and oiltemperature can change. At step 914, the counter i is increased by one,and the steps at 904-910 are repeated for the new engine speed andtransmission fluid characteristics. Steps 912, 914, 904, and 908 aresequentially repeated until it is determined at 910 that i=M. At thatpoint, it can be interpolated that a sufficient amount of data iscollected to construct a transfer function and create a map of the oilflow at clutch C1 for various engine speeds and oil flow rates. This isdone at step 914. The method concludes for the first clutch C1 at step916.

FIG. 10 illustrates an example of an oil flow map at clutch C1 at firstgear created from the test procedures accomplished according to the flowchart 900. As can be seen in this embodiment, the oil flow rate atclutch C1 increases as engine speeds increases, for both a lower oiltemperature and a higher oil temperature.

FIG. 11 is a flow chart 1100 of a method for generating and mapping oilflow at additional clutches Cn in a corresponding number n of differentgears. Steps 1102, 1104, 1108, 1110, 1112, 1114, and 1116 mirror thetest steps of FIG. 9, except that new tests are run for each of thedifferent clutches Cn and each of the number n of gears. Like FIG. 10for the first clutch C1, FIG. 12 illustrates an example of oil flow mapat another clutch Cn at an n-th gear.

FIG. 13A illustrates an oil flow map at the first clutch C1 at firstgear with a higher engine oil temperature and a lower oil temperature asengine speed is varied. The map is created based on the test methodsdescribed above. FIG. 13B shows a similar oil flow map, but for a secondclutch C2 at a second gear. FIG. 13C shows another oil flow map foranother clutch Cn for an n-th gear. Various maps can be created for acorresponding number of clutches and gears in the transmission.

The methods described above allow for identification of automatictransmission lubrication oil flow maps under actual drive conditionswithout having actual flow measurements during actual driving. Thisallows for accurate calibration of controls within the vehicle, as thecharacteristics oil can be estimated with a high degree of certaintybased on the test data.

The maps and models described above, and the test methods forformulating the maps, can be tuned and adjusted for various vehiclespecifications. For example, orifice sizes, tube lengths, and othersystem parameters can be varied for optimized data collection andanalysis.

The processes, methods, or algorithms disclosed herein can bedeliverable to/implemented by a processing device, controller, orcomputer, which can include any existing programmable electronic controlunit or dedicated electronic control unit. Similarly, the processes,methods, or algorithms can be stored as data and instructions executableby a controller or computer in many forms including, but not limited to,information permanently stored on non-writable storage media such as ROMdevices and information alterably stored on writeable storage media suchas floppy disks, magnetic tapes, CDs, RAM devices, and other magneticand optical media. The processes, methods, or algorithms can also beimplemented in a software executable object. Alternatively, theprocesses, methods, or algorithms can be embodied in whole or in partusing suitable hardware components, such as Application SpecificIntegrated Circuits (ASICs), Field-Programmable Gate Arrays (FPGAs),state machines, controllers or other hardware components or devices, ora combination of hardware, software and firmware components.

While exemplary embodiments are described above, it is not intended thatthese embodiments describe all possible forms encompassed by the claims.The words used in the specification are words of description rather thanlimitation, and it is understood that various changes can be madewithout departing from the spirit and scope of the disclosure. Aspreviously described, the features of various embodiments can becombined to form further embodiments of the invention that may not beexplicitly described or illustrated. While various embodiments couldhave been described as providing advantages or being preferred overother embodiments or prior art implementations with respect to one ormore desired characteristics, those of ordinary skill in the artrecognize that one or more features or characteristics can becompromised to achieve desired overall system attributes, which dependon the specific application and implementation. These attributes caninclude, but are not limited to cost, strength, durability, life cyclecost, marketability, appearance, packaging, size, serviceability,weight, manufacturability, ease of assembly, etc. As such, to the extentany embodiments are described as less desirable than other embodimentsor prior art implementations with respect to one or morecharacteristics, these embodiments are not outside the scope of thedisclosure and can be desirable for particular applications.

What is claimed is:
 1. A method of determining automatic transmissionlubrication fluid flow rates corresponding to a running vehicle withoutdirect oil flow measurements, the method comprising: obtaining a set ofin-vehicle clutch torques for a chosen clutch pack during a gear shiftevent for a set of shift conditions; performing a series of bench testsat various clutch pack clearances and lubrication oil flow rates for theset of shift conditions; adjusting a set of clutch pack clearances andlubrication oil flow rates during the series of bench tests in responseto a difference between bench-test measured clutch torques and thecorresponding in-vehicle clutch torques exceeding a threshold;estimating an in-vehicle transmission lubrication oil flow rate at thechosen clutch pack for the set of shift conditions in response to thedifference between bench-test measured clutch torques and thecorresponding in-vehicle clutch torques being less than the threshold;recording the estimated real in-vehicle transmission lubrication oilflow rate for a first engine condition and transmission oil temperature,corresponding to a given transmission gear position; reproducing thesteps of measuring, performing, adjusting, estimating, and recording forother engine conditions and transmission fluid temperatures,corresponding to other given transmission gear positions; constructing afunctional map of in-vehicle transmission lubrication oil flow rates atthe clutch packs for the given transmission gear positions for the firstand other engine conditions and transmission oil temperatures; andadjusting controls of the transmission based on the functional map ofin-vehicle transmission lubrication oil flow rates for desiredlubrication oil distributions within the transmission.
 2. The method ofclaim 1, wherein the engine conditions include engine speed during theshift condition.
 3. A method comprising: obtaining in-vehicle clutchtorques (IVCTs) for a clutch pack during a gear shift event in atransmission; performing bench tests at various lubrication oil-flowrates to obtain bench-test measured clutch torques (BTMCTs); adjustingthe oil-flow rates during the performing based on a difference betweenthe BTMCTs and the IVCTs; and controlling the transmission accountingfor an estimated in-vehicle lubrication oil-flow rate that is based onthe difference being less than a threshold.
 4. The method of claim 3,wherein the performing includes performing a series of bench tests atvarious clutch pack clearances and lubrication oil flow rates for a setof shift conditions.
 5. The method of claim 3, wherein the adjustingincludes adjusting a set of clutch pack clearances during the performingof bench tests in response to the difference between bench-test measuredclutch torques and the corresponding in-vehicle clutch torques exceedinga threshold.
 6. The method of claim 3, further comprising recording theestimated real in-vehicle lubrication oil-flow rate for a first enginecondition and a first transmission oil temperature, corresponding to agiven transmission gear position.
 7. The method of claim 6, furthercomprising reproducing the steps of obtaining, performing, adjusting,and recording for other engine conditions and other transmission fluidtemperatures, corresponding to other transmission gear positions.
 8. Themethod of claim 7, further comprising constructing a functional map ofin-vehicle transmission oil flow rates for the first and othertransmission gear positions, the first and other engine conditions, andthe first and other transmission oil temperatures.
 9. The method ofclaim 8, further comprising adjusting controls of the transmission basedon the functional map of in-vehicle transmission oil flow rates fordesired lubrication oil distributions within the transmission.